Vehicle control device and vehicle control method

ABSTRACT

A vehicle control device includes a state quantity detection device and a friction brake orientation control device. The state quantity detection device is configured to detect a state quantity indicating a vehicle body orientation. The friction brake orientation control device is configured to minimize pitching motion in the vehicle body orientation by applying braking torque from a friction brake at least to a front wheel, and minimize bouncing motion in the vehicle body orientation by applying braking torque from the friction brake to four wheels. The friction brake orientation control device is configured to prioritize minimizing the pitching motion over minimizing the bouncing motion.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application is a U.S. National stage application of InternationalApplication No. PCT/JP2013/051183, filed Jan. 22, 2013, which claimspriority to Japanese Patent Application No. 2012-012592 filed in Japanon Jan. 25, 2012, the contents of each of which are hereby incorporatedherein by reference.

BACKGROUND

1. Field of the Invention

The present invention relates to a control device for controlling thestate of a vehicle.

2. Background Information

Japanese Laid-Open Patent Application No. 2009-127456 discloses atechnology related to a vehicle control device. Said publicationdiscloses a technique of minimizing bouncing and pitching motion in avehicle body by controlling wheel torque.

SUMMARY

Minimizing bouncing and pitching motion using the friction brakes thatstop the vehicle will generate a strong sense of deceleration,potentially creating an unnatural sensation for a passenger.

The present invention was conceived in view of the problem describedabove, and has an object of providing a vehicle control device capableof controlling vehicle body orientation while reducing the amount ofunnatural ride sensation experienced by a passenger.

In order to achieve the abovementioned object, pitching motion in thevehicle body orientation is minimized in the vehicle control deviceaccording to the present invention by applying braking torque from afriction brake to at least a front wheel, and braking torque from thefriction brake is applied to four wheels to prioritize minimizingpitching motion over minimizing bouncing motion when performing frictionbrake orientation control for minimizing bouncing motion in the vehiclebody orientation.

Specifically, prioritizing minimizing pitching motion over minimizingbouncing motion when controlling vehicle body orientation using thefriction brakes allows the sensation of deceleration to be minimized,reducing the level of unnatural sensations experienced by a passenger.

BRIEF DESCRIPTIONS OF THE DRAWINGS

Referring now to the attached drawings which form a part of thisoriginal disclosure.

FIG. 1 is a schematic system diagram of a vehicle control deviceaccording to a first embodiment.

FIG. 2 is a control block diagram showing a configuration of controlperformed by the vehicle control device according to the firstembodiment.

FIG. 3 is a control block diagram showing a configuration of roll rateminimization control according to the first embodiment.

FIG. 4 is a time chart showing an envelope waveform formation processperformed in the roll rate minimization control of the first embodiment.

FIG. 5 is a control block diagram showing the configuration of a drivingstate estimator unit of the first embodiment.

FIG. 6 is a control block diagram showing the specifics of control in astroke speed calculator unit of the first embodiment.

FIG. 7 is a block diagram showing the configuration of a reference wheelspeed calculator unit of the first embodiment.

FIGS. 8A and 8B are schematic diagrams of a vehicle body vibrationmodel.

FIG. 9 is a control block diagram of actuator control amount calculationprocesses performed during pitch control in the first embodiment.

FIG. 10 is a control block diagram of brake pitch control in the firstembodiment.

FIG. 11 is a graph simultaneously showing a wheel speed frequencyprofile detected by a wheel speed sensor and a stroke frequency profilefrom a stroke sensor not installed in the present embodiment.

FIG. 12 is a control block diagram showing frequency-sensitive controlin sprung mass vibration damping control in the first embodiment.

FIG. 13 is a correlation graph showing human sensation profiles indifferent frequency regions.

FIG. 14 is a plot showing the relationship between the proportion ofvibration contamination and damping force in a float region in thefrequency-sensitive control of the first embodiment.

FIG. 15 is a wheel speed frequency profile detected by a wheel speedsensor in certain driving conditions.

FIG. 16 is a block diagram showing a control configuration for unsprungmass vibration damping control in the first embodiment.

FIG. 17 is a control block diagram showing a control configuration for adamping force control unit of the first embodiment.

FIG. 18 is a flow chart of a damping coefficient reconciliation processperformed during a standard mode in the first embodiment.

FIG. 19 is a flow chart of a damping coefficient reconciliation processperformed during a sports mode in the first embodiment.

FIG. 20 is a flow chart of a damping coefficient reconciliation processperformed during a comfort mode in the first embodiment.

FIG. 21 is a flow chart of a damping coefficient reconciliation processperformed during a highway mode in the first embodiment.

FIG. 22 is a time chart showing changes in damping coefficient whendriving on a hilly road surface and a bumpy road surface.

FIG. 23 is a flow chart of a driving state-based mode selection processperformed by a damping coefficient-reconciling unit of the firstembodiment.

DETAILED DESCRIPTION OF THE EMBODIMENTS First Embodiment

FIG. 1 is a schematic system diagram of a vehicle control deviceaccording to a first embodiment. A vehicle comprises an engine 1constituting a power source, brakes 20 for generating braking torque byapplying frictional force to the wheels (brakes corresponding toindividual wheels will be referred to hereafter as follows: front rightbrake: 20FR; front left brake: 20FL; rear right brake: 20RR; rear leftbrake: 20RL), and variable-damping-force shock absorbers 3 providedbetween each of the wheels and the vehicle body (“shock absorber” willbe abbreviated “S/A” in the following description; shock absorberscorresponding to individual wheels will be referred to as follows: frontright S/A: 3FR; front left S/A: 3FL; rear right S/A: 3RR; rear left S/A:3RL).

The engine 1 comprises an engine controller 1 a (also referred tohereafter as an engine control unit; equivalent to a motive power sourcecontrol device) for controlling the torque outputted by the engine 1;the engine controller 1 a controls the engine operation state (enginerpm, engine output torque, etc.) as desired by controlling the openingof the throttle valve, the fuel injection level, the ignition timing,and the like of the engine 1. The brakes 20 generate braking torque onthe basis of hydraulic pressure supplied from a brake control unit 2capable of controlling brake hydraulic pressure for each of the wheelsaccording to driving state. The brake control unit 2 comprises a brakecontroller 2 a (also referred to hereafter as a brake control unit) forcontrolling the braking torque generated by the brakes 20; the desiredhydraulic pressure is generated in the brakes 20 for each of the wheelsby the opening and closing of a plurality of solenoid valves usingmaster cylinder pressure generated by a driver operating the brake pedalor pump pressure generated by a built-in motor-driven pump as ahydraulic pressure source.

The S/As 3 are damping force-generating devices for damping the elasticmotion of coil springs provided between the unsprung mass (the axles,wheels, etc.) and the sprung mass (vehicle body, etc.) of the vehicle,and the damping force generated thereby can be adjusted by the operationof actuators. Each of the S/As 3 comprises a cylinder in which fluid issealed, a piston that makes strokes within the cylinder, and an orificefor controlling the movement of the fluid between fluid chambers formedabove and below the piston. Orifices of various diameters are formed inthe piston, and an orifice corresponding to a control command isselected from the various orifices when the S/A actuator operates.Damping force corresponding to the diameter of the orifice is therebygenerated. The movement of the piston will be more easily restricted ifthe orifice diameter is small, increasing damping force, and movement ofthe piston will be less easily restricted if the orifice diameter islarge, decreasing damping force.

Apart from selecting the diameter of the orifice, damping force may alsobe set, for example, by disposing a solenoid control valve over apassage linking the fluid chambers formed above and below the piston andcontrolling the opening and closing of the solenoid control valve; theinvention is not particularly limited with respect to this point. Eachof the S/As 3 comprises an S/A controller 3 a (equivalent to a dampingforce control device) for controlling the damping force of the S/A 3,and damping force is controlled by the operated of the orifice diameterby the S/A actuator.

Also comprised are wheel speed sensors 5 for detecting the wheel speedof each of the wheels (the sensors will be referred to as follows whenwheel speeds corresponding to individual wheels are indicated: frontright wheel speed: 5FR; front left wheel speed 5FL; rear right wheelspeed: 5RR; rear left wheel speed: 5RL), an integrated sensor 6 fordetecting forward/reverse acceleration, yaw rate, and lateralacceleration acting upon the center of gravity of the vehicle, asteering angle sensor 7 for detecting a steering angle indicating theamount to which the driver has operated the steering wheel, a vehiclespeed sensor 8 for detecting vehicle speed, an engine torque sensor 9for detecting engine torque, an engine rpm sensor 10 for detectingengine rpm, a master pressure sensor 11 for detecting master cylinderpressure, a brake switch 12 for outputting an on state signal when abrake pedal is operated, and an accelerator opening sensor 13 fordetecting the degree to which an accelerator pedal is open. Signals fromthe various sensors are inputted to the S/A controller 3 a. Theintegrated sensor 6 may be disposed at the center of gravity of thevehicle or at another location without restriction as long as the sensoris capable of estimating various values at the position of the center ofgravity. The sensor need not be integrated; individual sensors fordetecting yaw rate, forward/reverse acceleration, and lateralacceleration may also be provided.

FIG. 2 is a control block diagram showing a configuration of controlperformed by the vehicle control device according to the firstembodiment. The first embodiment comprises three controllers: the enginecontroller 1 a, the brake controller 2 a, and the S/A controller 3 a.The S/A controller 3 a comprises a driver input control unit 31 forperforming driver input control so that a desired vehicle orientation isattained on the basis of driver operations (of the steering wheel,accelerator and brake pedals, etc.), a driving state estimator unit 32for estimating driving state on the basis of values detected by thesensors, a sprung mass vibration damping control unit 33 for controllingthe vibrational state of the sprung mass of the vehicle on the basis ofthe estimated driving state, an unsprung mass vibration damping controlunit 34 for controlling the vibrational state of the unsprung mass ofthe vehicle on the basis of the estimated driving state, and a dampingforce control unit 35 for deciding upon the damping force to be set forthe S/As 3 on the basis of a shock absorber orientation control amountoutputted from the driver input control unit 31, a sprung mass vibrationdamping control amount outputted from the sprung mass vibration dampingcontrol unit 33, and an unsprung mass vibration damping control amountoutputted from the unsprung mass vibration damping control unit 34, andcontrolling S/A damping force.

In the first embodiment, three controllers are provided, but the presentinvention is not particularly limited to such a configuration; forexample, the damping force control unit 35 may be provided separatelyfrom the S/A controller 3 a as an orientation controller for a total offour S/A controllers including the damping force control unit 35, or allthe controllers may be combined into a single integrated controller. Theconfiguration of the first embodiment envisions the repurposing of anengine controller and a brake controller in an existing vehicle as anengine control unit 1 a and a brake control unit 2 a, and theinstallation of a separate S/A controller 3 a to create the vehiclecontrol device according to the first embodiment.

(Overall Configuration of Vehicle Control Device)

Three actuators are used in the vehicle control device according to thefirst embodiment in order to control the vibrational state of the sprungmass of the vehicle. Because the control performed by each of theactivators affects the state of the sprung mass of the vehicle,interference is a problem. In addition, the elements controllable by theengine 1, the elements controllable by the brakes 20, and the elementscontrollable by the S/As 3 all differ, and the matter of thecombinations in which these elements should be controlled is anotherproblem. For example, the brakes 20 are capable of controlling bouncingmotion and pitching motion, but controlling both at the same time willcreate a strong sense of deceleration and tend to create an unnaturalfeel for the driver. The S/As 3 are capable of absorbing rolling motion,bouncing motion, and pitching motion, but controlling all three of theseusing the S/As 3 will lead to increased manufacturing costs for the S/As3, and tends to increase damping force; this facilitates high-frequencyvibrational input from the road surface, also creating an unnatural feelfor the driver. In other words, a trade-off must be made in that controlperformed by the brakes 20 will not lead to worse high-frequencyvibration but will lead to an increased sense of deceleration, andcontrol performed by the S/As 3 will not create a sense of decelerationbut will lead to high-frequency vibrational input.

Thus, a control configuration has been adopted for the vehicle controldevice of the first embodiment in which a comprehensive assessment ismade of these problems in order to draw upon the respective advantagesof these control methods so as to complement the weaknesses of theother, thereby yielding a vehicle control device that is both economicaland offers superior vibration damping performance. To this end, thefollowing points were taken into consideration in the overallconstruction of the control system.

(1) The amount of control performed by the S/As 3 is minimized byprioritizing control performed by the engine 1 and the brakes 20.

(2) The only type of motion subjected to control by the brakes 20 ispitching motion, thereby eliminating the sense of deceleration producedfrom control by the brakes 20.

(3) The amount of control performed by the engine 1 and the brakes 20 isrestricted to less than the actually outputtable control amount, therebyreducing the burden placed upon the S/As 3 and minimizing the unnaturalfeel yielded by control performed by the engine 1 and the brakes 20.

(4) Skyhook control is performed by all of the actuators. This allowsskyhook control to be inexpensively performed using all of the wheelspeed sensors installed in the vehicle, without the use of a strokesensor, sprung mass vertical acceleration sensor, or the like, as isusually necessary to perform skyhook control.

(5) Scalar control (frequency-sensitive control) has been newlyintroduced in order to address high-frequency vibrational input, whichis difficult to address using skyhook control or other types of vectorcontrol, when the S/As 3 are performing sprung mass control.

(6) The control state manifested by the S/As 3 is selected, asappropriate, according to the driving state, thereby providing a controlstate suited to the driving conditions.

The foregoing has been a summary of the features of the control systemaccording to the embodiment as a whole. The specifics by which each ofthese individual features will be described in sequence hereafter.

(Driver Input Control Unit)

First, the driver input control unit will be described. The driver inputcontrol unit 31 comprises an engine driver input control unit 31 a forattaining the vehicle orientation demanded by the driver by controllingthe torque of the engine 1, and an S/A driver input control unit 31 bfor attaining the vehicle orientation demanded by the driver bycontrolling the damping force of the S/As 3. The engine driver inputcontrol unit 31 a calculates a ground load variation minimizationcontrol amount for minimizing variations in the ground loads of thefront wheels and the rear wheels, and a yaw response control amountcorresponding to the vehicle behavior desired by the driver on the basisof signals from the steering angle sensor 7 and the vehicle speed sensor8.

The S/A driver input control unit 31 b calculates a driver input dampingforce control amount corresponding to the vehicle behavior desired bythe driver on the basis of the signals from the steering angle sensor 7and the vehicle speed sensor 8, and outputs this amount to the dampingforce control unit 35. If, for example, the nose of the vehicle riseswhile the driver is turning, the driver's field of view can easily betaken off the road; in such cases, the damping force of the four wheelsis outputted as the driver input damping force control amount so as toprevent the nose from rising. A driver input damping force controlamount for minimizing rolling generating during turning is alsooutputted.

(Controlling Rolling via S/A Driver Input Control)

Roll minimization control performed via S/A driver input control willnow be described. FIG. 3 is a control block diagram showing aconfiguration of roll rate minimization control according to the firstembodiment. A lateral acceleration estimator unit 31 b 1 estimateslateral acceleration Yg on the basis of a front wheel steering angle δfdetected by the steering angle sensor 7, a rear wheel steering angle δr(the actual rear wheel steering angle if a rear wheel steering device isprovided; otherwise, 0 as appropriate), and a vehicle speed VSP detectedby the vehicle speed sensor 8. The lateral acceleration Yg is calculatedaccording to the following formula using an estimated yaw rate value γ.

Yg=VSp·γ

The estimated yaw rate value γ is calculated according to the followingformula:

$\begin{Bmatrix}\beta \\\gamma\end{Bmatrix} = {N\begin{Bmatrix}\delta_{f} \\\delta_{r}\end{Bmatrix}}$ $\begin{Bmatrix}\beta \\\gamma\end{Bmatrix} = {M^{- 1}N\begin{Bmatrix}\delta_{f} \\\delta_{r}\end{Bmatrix}}$ wherein ${M = \begin{bmatrix}m_{11} & m_{12} \\m_{21} & m_{22}\end{bmatrix}},{N = \begin{bmatrix}n_{11} & n_{12} \\n_{21} & n_{22}\end{bmatrix}}$ m₁₁ = −(Ktf ⋅ Lf − Ktv ⋅ Lv)$m_{12} = {{- \frac{1}{V}}\left( {{{Ktf} \cdot {Lf}^{2}} - {{Ktv} \cdot {Lv}^{2}}} \right)}$m₂₁ = −2(Ktf + Ktv)$m_{22} = {{{- \frac{2}{V}}\left( {{{Ktf} \cdot {Lf}} - {{Ktv} \cdot {Lv}}} \right)} - {M \cdot V}}$n₁₁ = −Ktf ⋅ Lf n₁₂ = Ktv ⋅ Lr n₂₁ = −2 ⋅ Ktf n₂₂ = −2 ⋅ Ktv

β: vehicle body slide angle

γ: vehicle body yaw rate

δf: front wheel steering angle

δr: rear wheel steering angle

V: vehicle body

Ktf: front wheel CP

Ktv: rear wheel CP

Lf: distance between front axle and center of gravity

Lr: distance between rear axle and center of gravity

M: vehicle body mass

A 90° phase lead component-generating unit 31 b 2 differentiates theestimated lateral acceleration Yg and outputs a lateral accelerationderivative dYg. A 90° phase lag component-generating unit 31 b 3 outputsa component F(dYg) in which the phase of the lateral accelerationderivative dYg is delayed 90°. In the component F(dYg) 90°, the phase ofthe component from which the low-frequency region has been removedgenerated by the phase lead component-generated unit 31 b 2 is returnedto the phase of the lateral acceleration Yg, and the DC is cut from thelateral acceleration Yg; i.e., the component is a transitional componentof lateral acceleration Yg. A 90° phase lag component-generating unit 31b 4 outputs a component F(Yg) in which the phase of the lateralacceleration Yg is delayed 90°.

A gain multiplier unit 31 b 5 multiplies the lateral acceleration Yg,lateral acceleration derivative dYg, lateral acceleration DC-cutcomponent F(dYg), and 90° phase lag component F(Yg) by gain. Gain is setaccording to the roll rate transfer function for the steering angle.Gain may also be adjusted according to four control modes describedhereafter. A square calculator unit 31 b 6 squares and outputs thecomponents having been multiplied by gain. A synthesizer unit 31 b 7adds the values outputted by the squaring processor unit 31 b 6. A gainmultiplier unit 31 b 8 multiplies the square of the summed components bygain, and outputs the results. A square root calculator unit 31 b 9calculates the square root of the value outputted by the gain multiplierunit 31 b 7, thereby calculating a driver input orientation controlamount for use in roll rate minimization control, and outputs the amountto the damping force control unit 35.

The 90° phase lead component-generating unit 31 b 2, 90° phase lagcomponent-generating unit 31 b 3, 90° phase lag component-generatingunit 31 b 4, gain multiplier unit 31 b 5, square calculator unit 31 b 6,synthesizer unit 31 b 7, gain multiplier unit 31 b 8, and square rootcalculator unit 31 b 9 constitute a Hilbert transform unit 31 b 10 forgenerating an envelope waveform using a Hilbert transform.

FIG. 4 is a time chart showing an envelope waveform formation processperformed in the roll rate minimization control of the first embodiment.

When a driver begins steering at time t1, a roll rate gradually beginsto be generated. At this point, the 90° phase lead component dYg isadded to form an envelope waveform, and the driver input orientationcontrol amount is calculated on the basis of the envelope waveform-basedscalar quantity, thereby allowing roll rate generation during theinitial stage of steering to be minimized. In addition, the addition ofthe lateral acceleration DC-cut component F(dYg) to form the envelopewaveform allows roll rates generated during transitional states, such aswhen the driver is beginning or ending steering, to be efficientlyminimized. In other words, damping force is not excessively increasedand degradation in ride comfort can be avoided during constant turningstates in which roll is stably generated.

Next, when the driver holds the steering wheel in place at time t2, the90° phase lead component dYG and the lateral acceleration DC-cutcomponent F(dYg) disappear, and the 90° phase lag component F(Yg) isthen added. At this time, a roll rate resonance component equivalent tothe roll aftershock is generated after rolling occurs, even if there islittle change in the roll rate itself in a steady steering state. If thephase lag component F(Yg) were not added, damping force would be set toa low value from time t2 to time t3, risking destabilization of vehiclebehavior by the roll rate resonance component. The 90° phase lagcomponent F(Yg) is added in order to minimize the roll rate resonancecomponent.

When the driver shifts from the held steering state back to astraight-ahead driving state at time t3, the lateral acceleration Ygdecreases, and the roll rate is reduced to a low value. The action ofthe 90° phase lag component F(Yg) also ensures damping force at thispoint as well, allowing destabilization due to the roll rate resonancecomponent to be avoided.

(Driving State Estimator Unit)

Next, the driving state estimator unit will be described. FIG. 5 is acontrol block diagram showing the configuration of a driving stateestimator unit of the first embodiment. The driving state estimator unit32 of the first embodiment calculates a stroke speed, bounce rate, rollrate, and pitch rate for each wheel used in the skyhook controlperformed by the sprung mass vibration damping control unit 33 asdescribed hereafter primarily on the basis of the wheel speeds detectedby the wheel speed sensors 5. The values from the wheel speed sensors 5of the wheels are inputted into a stroke speed calculator unit 321, andsprung mass speed is calculated by the stroke speed calculator unit 321from the stroke speeds calculated for the wheels.

FIG. 6 is a control block diagram showing the specifics of control in astroke speed calculator unit of the first embodiment. A stroke speedcalculator unit 321 is separately provided for each wheel; the controlblock diagram shown in FIG. 6 focuses on a specific wheel. The strokespeed calculator unit 321 comprises a reference wheel speed calculatorunit 300 for calculating a reference wheel speed on the basis of thevalues from the wheel speed sensors 5, the front wheel steering angle δfdetected by the steering angle sensor 7, a rear wheel steering angle δr(the actual rear wheel steering angle if a rear wheel steering device isprovided, 0 otherwise), a vehicle body lateral speed, and an actual yawrate detected by the integrated sensor 6, a tire rotational vibrationfrequency calculator unit 321 a for calculating tire rotationalvibration frequency on the basis of the calculated reference wheelspeed, a deviation calculator unit 321 b for calculating the deviationbetween the reference wheel speed and the value from the wheel speedsensor (i.e., wheel speed variation), a GEO conversion unit 321 c forconverting the deviation calculated by the deviation calculator unit 321b to a suspension stroke amount, a stroke speed calibrator unit 321 dfor calibrating the converted stroke amount to a stroke speed, and asignal processing unit 321 e for applying a band elimination filtercorresponding to the frequency calculated by the tire rotationalvibration frequency calculator unit 321 a to the calibrated valueyielded by the stroke speed calibrator unit 321 d to eliminate a primarytire rotational vibration component and calculate a final stroke speed.

(Reference Wheel Speed Calculator Unit)

The reference wheel speed calculator unit 300 will now be described.FIG. 7 is a block diagram showing the configuration of a reference wheelspeed calculator unit of the first embodiment. The reference wheel speedis a wheel speed from which various types of interference from theindividual wheels have been removed. In other words, the differencebetween the value from the wheel speed sensor and the reference wheelspeed is related to a component that varies according to a strokegenerated by vehicle body bouncing motion, rolling motion, pitchingmotion, or unsprung vertical vibration; in the present embodiment, thestroke speed is calculated on the basis of this difference.

A flat surface motion component extractor unit 301 uses the wheel speedsensor values as inputs to calculate a first wheel speed V0 as areference wheel speed for each of the wheels on the basis of the vehiclebody plan view model. ω (rad/s) is the wheel speed sensor detected bythe wheel speed sensor 5, δf (rad) is a front wheel actual steeringangle detected by the steering angle sensor 7, δr (rad) is a rear wheelactual steering angle, Vx is vehicle body lateral speed, γ (rad/s) isthe yaw rate detected by the integrated sensor 6, V (m/s) is a vehiclebody speed estimated from the calculated reference wheel speed ω0, VFL,VFR, VRL, and VRR are the reference wheel speeds to be calculated, Tf isa front wheel treat, Tr is a rear wheel treat, Lf is the distance fromthe position of the vehicle center of gravity to the front wheels, andLr is the distance from the position of the vehicle center of gravity tothe rear wheel. The vehicle body plan view model is expressed as followsusing the symbols described above.

VFL=(V−Tf/2·γ)cos δf+(Vx+Lf·γ)sin δf

VFR=(V+Tf/2·γ)cos δf+(Vx+Lf·γ)sin δf

VRL=(V−Tf/2·γ)cos δr+(Vx+Lr·γ)sin δr

VRR=(V+Tf/2·γ)cos δr+(Vx+Lr·γ)sin δr   (Formula 1)

If a normal driving state in which no lateral sliding of the vehicleoccurs is hypothesized, 0 may be inputted for the vehicle body lateralspeed Vx. This yields the following formulas when the various formulasare rewritten with values based on V. When rewriting in this manner, Vis written as V0FL, V0FR, V0RL, and V0RR (equivalent to first wheelspeeds) as values corresponding to the various wheels.

V0FL={VFL−Lf·γ sin δf}/cos δf+Tf/2·γ

V0FR={VFR−Lf·γ sin δf}/cos δf−Tf/2·γ

V0RL={VRL+Lr·γ sin δr}/cos δr+Tr/2·γ

V0RR={VRR+Lf·γ sin δf}/cos δR+Tr/2·γ  (Formula 2)

A roll interference-removing unit 302 uses the first wheel speed V0 asan input to calculate second wheel speeds V0F, V0R as reference wheelspeeds for the front and rear wheels on the basis of a vehicle bodyfront view model. The vehicle body front view model is used to removewheel speed differences generated by rolling motion occurring around acenter of roll rotation on a normal line passing through the vehiclecenter of gravity as viewed from the front of the vehicle, and isrepresented as follows.

V0F=(V0FL+V0FR)/2

V0R=(V0RL+V0RR)/2

This yields second wheel speeds V0F, V0R from which roll-basedinterference has been removed.

A pitch interference-removing unit 303 uses the second wheel speeds V0F,V0R as inputs to calculate third wheel speeds VbFL, VbFR, VbRL, and VbRRconstituting reference wheel speeds for all the wheels according to avehicle body side view model. The vehicle body side view model is usedto remove wheel speed differences generated by pitching motion occurringaround a center of pitch rotation on a normal line passing through thevehicle center of gravity as viewed from the side of the vehicle.

VbFL=VbFR=VbRL=VbRR={Lr/(Lf+Lr)}V0F+{Lf/(Lf+Lr)}V0R   (Formula 3)

A reference wheel speed redistribution unit 304 uses VbFL(=VbFR=VbRL=VbRR) for V in the vehicle body plan view model shown informula 1 to calculate final reference wheel speeds VFL, VFR, VRL, VRRfor each wheel, which are divided by the tire radius r0 to calculate thereference wheel speed ω0.

Once the reference wheel speed ω0 has been calculated according to theprocess described above, the deviation between the reference wheel speedω0 and the wheel speed sensor value is calculated; the deviationrepresents a wheel speed variation arising from suspension strokes, andis therefor converted into a stroke speed Vzs. As a rule, not only doesa suspension make strokes in the vertical direction when holding thewheels, but the wheel rotational centers move forwards and backwards asstrokes occur, and the axles equipped with the wheel speed sensors 5become tilted, creating a difference in rotational angle with thewheels. Because this forward and backward motion leads to changes inwheel speed, deviations between the reference wheel speed and the wheelspeed sensor value can be extracted as stroke-induced variations. Thedegree of variation that occurs can be set, as appropriate, according tothe suspension geometry.

Once the stroke speed calculator unit 321 has calculated the strokespeeds Vz_sFL, Vz_sFR, Vz_sRL, Vz_sRR for each wheel according to theprocess described above, a sprung mass speed calculator unit 322calculates a bounce rate, roll rate, and pitch rate for use in skyhookcontrol.

(Estimation Model)

In skyhook control, damping force is set according to the relationshipbetween the stroke speeds of the S/As 3 and the sprung mass speed, andthe orientation of the sprung mass is controlled to achieve a flatdriving state. In order to achieve control of the orientation of thesprung mass via skyhook control, feedback on the sprung mass speed isnecessary. Stroke speed is a value detectable from the wheel speedsensor 5; since a sensor for the vertical acceleration of the sprungmass is not provided, the sprung mass speed must be estimated using anestimation model. Problems involved in the estimation model and theappropriate model configuration to adopt will now be discussed.

FIGS. 8A and 8B are schematic diagrams of a vehicle body vibrationmodel. FIG. 8A is a model for a vehicle provided with S/As of constantdamping force (hereafter referred to as a conventional vehicle), andFIG. 8B is a model for a vehicle provided with variable S/As in whichskyhook control is performed. In FIGS. 8A and 8B, Ms indicates sprungmass, Mu indicates unsprung mass, Ks indicates coil spring modulus ofelasticity, Cs indicates S/A damping coefficient, Ku indicates unsprung(tire) modulus of elasticity, Cu indicates unsprung (tire) dampingcoefficient, and Cv indicates a variable damping coefficient. z2indicates the position of the sprung mass, z1 indicates the position ofthe unsprung mass, and z0 indicates the position of the road surface.

If the conventional vehicle model shown in FIG. 8A is used, the equationof motion for the sprung mass is expressed as follows. The first-orderdifferential for z1 (i.e., speed) is represented by dz1, and thesecond-order differential (i.e., acceleration) is represented by ddz1.

Ms·ddz2=−Ks(z2−z1)−Cs(dz2−dz1)   (Estimation Formula 1)

Applying a Laplace transform to this relationship yields the followingformula.

dz2=−(1/Ms)·(1/S ²)·(Cs·s+Ks)(dz2−dz1)   (Estimation Formula 2)

Because dz2−dz1 is stroke speed (Vz_sFL, Vz_sFR, Vz_sRL, Vz_sRR), thesprung mass speed can be calculated from the stroke speed. However,adjusting damping force via skyhook control will vastly reduceestimation precision, creating the problem that a large orientationcontrol force (damping force adjustment) cannot be applied in theconventional vehicle model.

Thus, the use of a skyhook control-based vehicle model shown in FIG. 8Bis conceivable. As a rule, altering damping force involves altering theforce limiting the piston movement speed of the S/As 3 as suspensionstrokes occur. Because semi-active S/As 3 in which the pistons cannot beactively moved in a desired direction, a semi-active skyhook model isused; calculating sprung mass speed yields the following formula.

dz2=−(1/Ms)·(1/s ²)·{(CS+CV)·s+Ks}(dz2−dz1)   (Estimation Formula 3)

wherein:

if dz2·(dz2−dz1)≧0, Cv=Csky·{dz21(dz2−dz1)}, and

if dz2·(dz2−dz1)<0, Cv=0.

That is, Cv is a discontinuous value.

If the semi-active skyhook model is viewed as a filter when one wishesto estimate sprung mass speed using a simple filter, the variables areequivalent to filter coefficients, and a variable damping coefficient Cvthat is discontinuous in the pseudo-differential term {(Cs+Cv)·s+Ks};thus, filter response is unstable, and suitable estimation precisioncannot be obtained. In particular, unstable filter response will lead tophase shifting. Skyhook control cannot be achieved if the correspondencebetween the phase and the sign for sprung mass speed breaks down. Thedecision was thus made to estimate sprung mass speed using an activeskyhook model, in which a stable Csky can be directly used withoutrelying upon the sign relationship between sprung mass speed and thestroke speed even if semi-active S/As 3 are used. The use of an activeskyhook model to calculate sprung mass speed can be expressed asfollows.

dz2=−(1/s)·{1/(s+Csky/Ms)}·{(Cs/Ms)s+(Ks/Ms)}(dz2−dz1)   (EstimationFormula 4)

In this case, there is no discontinuity in the pseudo-differential term{(Cs/Ms)s+(Ks/Ms)}, and the term {1/(s+Csky/Ms)} can be constitutedusing a low-pass filter. Filter response is therefore stable, andsuitable estimation precision is obtainable. It should be noted that,even if an active skyhook model is adopted, only semi-active control isactually possible; thus, the controllable range is halved. The estimatedsprung mass speed is therefore less than the actual speed in thefrequency band from sprung mass resonance down. However, it is the phasethat is most vital in the context of skyhook control, and skyhookcontrol can be achieved as long as the correspondence between phase andsign can be maintained; this is unproblematic because the sprung massspeed can be adjusted using the other coefficients or the like.

It is apparent from the relationship described above that sprung massspeed can be estimated if the stroke speeds for each wheel are known.Because an actual vehicle has not one wheel, but four, we will nowconsider using the stroke speeds for each wheel to estimate the state ofthe sprung mass divided into roll rate, pitch rate, and bounce ratemodes. When calculating the abovementioned three components from thestroke speeds of the four wheels, one corresponding component islacking, leading to an indefinite solution; thus, warp rate, whichindicates the movement of diagonally opposed wheels, has beenintroduced. Defining xsB as the bounce term of the stroke amount, xsR asthe roll term, xsP as the pitch term, xsW as the warp term, and z_sFL,z_sFR, z_sRL, z_sRR as stroke amounts corresponding to Vz_sFL, Vz_sFR,Vz_sRL, Vz_sRR, the following formula arises.

$\begin{matrix}{\begin{Bmatrix}{z\_ sFL} \\{z\_ sFR} \\{z\_ sRL} \\{z\_ sRR}\end{Bmatrix} = {{\begin{bmatrix}1 & 1 & {- 1} & {- 1} \\1 & {- 1} & {- 1} & 1 \\1 & 1 & 1 & 1 \\1 & {- 1} & 1 & {- 1}\end{bmatrix}\begin{Bmatrix}{xsB} \\{xsR} \\{xsP} \\{xsW}\end{Bmatrix}\begin{Bmatrix}{xsB} \\{xsR} \\{xsP} \\{xsW}\end{Bmatrix}} = {\begin{bmatrix}1 & 1 & {- 1} & {- 1} \\1 & {- 1} & {- 1} & 1 \\1 & 1 & 1 & 1 \\1 & {- 1} & 1 & {- 1}\end{bmatrix}^{- 1}\begin{Bmatrix}{z\_ sFL} \\{z\_ sFR} \\{z\_ sRL} \\{z\_ sRR}\end{Bmatrix}}}} & \left( {{Formula}\mspace{14mu} 1} \right)\end{matrix}$

In view of the relationship shown above, the differentials dxsB, . . .of xsB, xsR, xsP, xsW are expressed by the following formulas.

dxsB=1/4(Vz _(—) sFL+Vz _(—) sFR+Vz _(—) sRL+Vz _(—) sRR)

dxsR=1/4(Vz _(—) sFL−Vz _(—) sFR+Vz _(—) sRL−Vz _(—) sRR)

dxsP=1/4(−Vz _(—) sFL−Vz _(—) sFR+Vz _(—) sRL+Vz _(—) sRR)

dxsW=1/4(−Vz _(—) sFL+Vz _(—) sFR+Vz _(—) sRL−Vz _(—) sRR)

The relationship between sprung mass speed and stroke speed is obtainedfrom estimation formula 4 above; thus, taking G as the section ofestimation formula 4 reading −(1/s)·{1/(s+Csky/Ms)}·{(Cs/Ms)s+(Ks/Ms)},GB, GR, and GP as values taking into account modal parameters (CskyB,CskyR, CskyP, CsB, CsR, CsP, KsB, KsR, KsP) for the bounce terms, rollterms, and pitch terms of Csky, Cs, and Ks, respectively, dB as bouncerate, dR as roll rate, and dP as pitch rate, dB, dR, and dP can becalculated as follows.

dB=GB·dxsB

dR=GR·dxsR

dP=GP·dxsP

As shown from the foregoing, the state of the sprung mass of an actualvehicle can be estimated on the basis of the stroke speeds for thevarious wheels.

(Sprung Mass Vibration Damping Control Unit)

Next, the configuration of the sprung mass vibration damping controlunit 33 will be described. As shown in FIG. 2, the sprung mass vibrationdamping control unit 33 comprises a skyhook control unit 33 a forcontrolling orientation according to the estimated value for sprung massspeed described above, and a frequency-sensitive control unit 33 b forminimizing sprung mass vibration on the basis of the road surface inputfrequency.

(Configuration of Skyhook Control Unit)

The vehicle control device according to the first embodiment comprisesthree actuators for achieving sprung mass orientation control in theform of the engine 1, the brakes 20, and the S/As 3. Of these, bouncerate, roll rate, and pitch rate are the objects of control for the S/As3, bounce rate and pitch rate are the objects of control for the engine1, and pitch rate is the object of control for the brakes 20 in theskyhook control unit 33 a. In order to allocate control amounts to aplurality of actuators that act in different manners and control thestate of the sprung mass, a shared control amount must be used for each.In the first embodiment, the control amount for each of the actuatorscan be determined by using the sprung mass speed estimated by thedriving state estimator unit 32.

Bounce-directional skyhook control amount:

FB=CskyB·dB

Roll-directional skyhook control amount:

FR=CskyR·dR

Pitch-directional skyhook control amount:

FP=CskyP·dP

FB is sent to the engine 1 and the S/As 3 as the bounce orientationcontrol amount. FR is for control performed only by the S/As 3, and sois sent to the damping force control unit 35 as a roll orientationcontrol amount.

Next, the pitch-directional skyhook control amount FP will be described.Pitch control is performed by the engine 1, the brakes 20, and the S/As3.

FIG. 9 is a control block diagram of actuator control amount calculationprocesses performed during pitch control in the first embodiment. Theskyhook control unit 33 a comprises a first target orientation controlamount calculator unit 331 for calculating a target pitch rateconstituting a first target orientation control amount that is a controlamount that can used in common for all of the actuators, an engineorientation control amount calculator unit 332 for calculating theengine orientation control amount achieved by the engine 1, a brakeorientation control amount calculator unit 334 for calculating the brakeorientation control amount achieved by the brakes 20, and an S/Aorientation control amount calculator unit 336 for calculating the S/Aorientation control amount achieved by the S/As 3.

As the skyhook control according to the present system gives foremostpriority to pitch rate minimization, the first target orientationcontrol amount calculator unit 331 outputs pitch rate without furthermodification (hereafter, this pitch rate will be referred to as thefirst target orientation control amount). The engine orientation controlamount calculator unit 332 calculates the engine orientation controlamount, which is the control amount achievable by the engine 1, on thebasis of the inputted first target orientation control amount.

A limit value that limits the engine torque control amount according tothe engine orientation control amount so as not to create an unnaturalsensation for the driver is set in the engine orientation control amountcalculator unit 332. The engine torque control amount is thus keptwithin a predetermined forward/reverse acceleration range when convertedto forward/reverse acceleration. Accordingly, when the engine torquecontrol amount is calculated on the basis of the first targetorientation control amount and results in a value that is equal to orgreater than the limit value, a skyhook control amount for the pitchrate attainable using the limit value (i.e., the product of the pitchrate minimized by the engine 1 multiplied by CskyP; hereafter, “engineorientation control amount”) is outputted. At this time, a valueconverted to the pitch rate by a converter unit 332 a is outputted to asecond target orientation control amount calculator unit 333, to bedescribed hereafter. In addition, the engine control unit 1 a calculatesthe engine torque control amount on the basis of the engine orientationcontrol amount corresponding to the limit value, and outputs this amountto the engine 1.

The second target orientation control amount calculator unit 333calculates the difference between the first target orientation controlamount and the value of the engine orientation control amount convertedto pitch rate by the converter unit 332 a, and outputs this differenceto the brake orientation control amount calculator unit 334. A limitvalue for limiting the braking torque control amount so as not to createan unnatural feel for the driver, as in the case of the engine 334, isset in the brake orientation control amount calculator unit 334 (thislimit value will be discussed in detail hereafter).

The braking torque control amount is thus kept within a predeterminedforward/reverse acceleration range when converted to forward/reverseacceleration (with the limit value being calculated on the basis ofnaturalness of ride feel for passengers, actuator lifespan, etc.).Accordingly, when the brake orientation control amount is calculatedbased on the second target orientation control amount, and thecalculated value is equal to or greater than the limit value, a pitchrate minimization amount (hereafter, “brake orientation control amount”)achievable by such a limit value is outputted. At this time, a valueconverted to the pitch rate by a converter unit 3344 is outputted to athird target orientation control amount calculator unit 335, to bedescribed hereafter. In addition, the brake control unit 2 a calculatesa braking torque control amount (or deceleration) on the basis of thebrake orientation control amount corresponding to the limit value, andis outputted to the brake control unit 2.

The third target orientation control amount calculator unit 335calculates a third target orientation control amount constituted by thedifference between the second target orientation control amount and thebrake orientation control amount, and outputs this amount to the S/Aorientation control amount calculator unit 336. The S/A orientationcontrol amount calculator unit 336 outputs a pitch orientation controlamount corresponding to the third target orientation control amount.

The damping force control unit 35 calculates the damping force controlamount on the basis of the bounce orientation control amount, the rollorientation control amount, and the pitch orientation control amount(hereafter collectively referred to as the S/A orientation controlamounts), and outputs this amount to the S/As 3.

(Brake Pitch Control)

Brake pitch control will now be described. Generally, the brakes 20 arecapable of controlling both bounce and pitch; thus, it is preferablethat they control both. However, when bounce control is performed by thebrakes 20, braking force is applied to all four wheels simultaneously,and there is a strong sense of deceleration even in directions of lowcontrol priority despite the difficulty in obtaining control effects,tending to create an unnatural feed for the driver. Thus, aconfiguration in which the brakes 20 specialize in pitch control hasbeen adopted. FIG. 10 is a control block diagram of brake pitch controlin the first embodiment. Defining m as the mass of the vehicle body, Bffas front wheel braking force, BFr as rear wheel braking force, Hcg asthe height between the vehicle center of gravity and the road surface, aas vehicle acceleration, Mp as pitch moment, and Vp as pitch rate, thefollowing relationships hold.

BFf+BFr=m·a

m·a·Hcg=Mg

Mp=(BFf+BFr)·Hcg

If braking force is applied when the pitch rate Vp is positive, i.e.,the front wheel side of the vehicle is lowered, the front wheel sidewill sink further lower, augmenting pitch motion; thus, braking force isnot applied in such cases. On the other hand, when the pitch rate Vp isnegative, i.e., the front wheel side of the vehicle is raised, thebraking pitch moment will impart braking force, minimizing the rising ofthe front wheel side. This ensures the driver's field of view and makesthe area ahead easier to see, contributing to improved senses of safetyand flatness of ride. In other words, the following control amounts areapplied:

when Vp>0 (front wheels lowered), Mp=0; and

when Vp≦0 (front wheels raised), Mp=CskyP·Vp

Braking torque is thus generated only when the front side of the vehicleis raised, thereby allowing the sense of deceleration created thereby tobe reduced compared to cases in which braking torque is generated bothwhen the front side is raised and when it is lowered. In addition, theactuators need only be operated at half the frequency as usual, allowinginexpensive actuators to be used.

The brake orientation control amount calculator unit 334 comprises thefollowing control blocks on the basis of the relationship describedabove. A dead band process sign determiner unit 3341 determines the signfor the inputted pitch rate Vp; if the sign is positive, no control isnecessary, so 0 is outputted to a deceleration sense-reducing processor3342, and if the sign is negative, control is determined to be possible,so a pitch rate signal is outputted to the deceleration sense-reducingprocessor 3342.

(Deceleration Sense Reduction Process)

Next, a deceleration sense reduction process will be described. Thisprocess corresponds to the limit applied by the limit value set in thebrake orientation control amount calculator unit 334. A squaringprocessor 3342 a squares the pitch rate signal. This reverses the sign,and smoothes the increase in control force. A pitch rate square dampingmoment calculator unit 3342 b multiplies the squared pitch rate by askyhook gain CskyP for the pitch term that takes the squaring processinto account to calculate pitch moment Mp. A target decelerationcalculator unit 3342 c divides the pitch moment Mp by mass m and theHeight Hcg between the vehicle center of gravity and the road surface tocalculate a target deceleration.

A jerk threshold value limiter unit 3342 d determines whether the rateof change in the calculated target deceleration, i.e., jerk, is withinpreset deceleration jerk threshold value and release jerk thresholdvalue ranges, and whether the target deceleration is within aforward/reverse acceleration limit value range. If any of the thresholdvalues are exceeded, the target deceleration is corrected to a valuewithin the ranges for the jerk threshold values. If the targetdeceleration exceeds the limit value, it is set to within the limitvalue. It is thereby possible to generate deceleration so as not tocreate an unnatural feel for the driver.

A target pitch moment converter unit 3343 multiplies the targetdeceleration limited by the jerk threshold value limiter unit 3342 d bythe mass m and the height Hc9 to calculate a target pitch moment, whichis outputted to the brake control unit 2 a and a target pitch rateconverter unit 3344. The target pitch rate converter unit 3344 dividesthe target pitch moment by the pitch term skyhook gain CskyP to convertto a target pitch rate (equivalent to a brake orientation controlamount), which is outputted to a third target orientation control amountcalculator unit 335.

For pitch rate, as described above, the first target orientation controlamount is calculated, the engine orientation control amount is nextcalculated, the brake orientation control amount is calculated from thesecond target orientation control amount constituted by the differencebetween the first target orientation control amount and the engineorientation control amount, and the S/A orientation control amount iscalculated from the third target orientation control amount constitutedby the difference between the second orientation control amount and thebrake orientation control amount. It is thereby possible to reduce theamount of pitch rate control performed by the S/As 3 by the controlperformed by the engine 1 and the brakes 20, thus allowing for arelative reduction in the controllable range of the S/As 3, and sprungmass orientation control to be achieved using inexpensive S/As 3.

An increase in the amount of control performed by the S/As 3 will, ingeneral, increase damping force. Because increased damping force leadsto stiff suspension characteristics, high-frequency input is more easilytransmitted when there is high-frequency vibration input from the roadsurface, reducing passenger comfort (this situation is hereafterreferred to as exacerbated high-frequency vibration characteristics). Bycontrast, controlling pitch rate using the engine 1 and the brakes 20,which are actuators whose vibration transmission profiles are notaffected by road surface inputs, and reducing the amount of controlperformed by the S/As 3 allows the exacerbation of high-frequencyvibration characteristics to be avoided. The effects described above areobtained by determining the amount of control performed by the engine 1before that performed by the S/As 3, and the amount of control performedby the brakes 20 before that performed by the S/As 3.

(Frequency-Sensitive Control Unit)

Next, a frequency-sensitive control process performed in the sprung massvibration damping control unit will be described. In the firstembodiment, as a rule, the sprung mass speed is estimated on the basisof the values detected by the wheel speed sensors 5, and skyhook controlis performed based thereon, thereby achieving sprung mass vibrationdamping control. However, there are cases in which it may not bepossible to guarantee sufficient estimation precision using the wheelspeed sensors 5, and cases in which it is desirable to activelyguarantee a comfortable driving state (i.e., a soft ride rather than afeeling of vehicle body flatness) depending on driving conditions or thedriver's intent. In such cases, it may be difficult to effect suitablecontrol due to slight phase shifts if vector control such as skyhookcontrol, in which the relationship (phase, etc.) of the signs of thestroke speed and the sprung mass speed is vital, is used; thus,frequency-sensitive control constituted by sprung mass vibration dampingcontrol according to vibration profile scalar quantities has beenintroduced.

FIG. 11 is a graph simultaneously showing a wheel speed frequencyprofile detected by a wheel speed sensor and a stroke frequency profilefrom a stroke sensor not installed in the present embodiment. In thiscontext, “frequency profile” refers to a profile in which the magnitudeof amplitude against frequency is plotted on the y axis. A comparison ofthe frequency component of the wheel speed sensor 5 and the frequencycomponent of the stroke sensor shows that roughly similar scalarquantities can be plotted from the sprung mass resonance frequencycomponent to the unsprung mass resonance frequency component. Thus, thedamping force has been set on the basis of this frequency profile out ofthe values detected by the wheel speed sensor 5. The region in which thesprung mass resonance frequency component is present is a frequencyregion in which the swaying of a passenger's entire body creates asensation as thought the passenger is floating in the air, that is, thatthe gravitational acceleration affecting the passenger has decreased,and is referred to as the float region (0.5-3 Hz). The region betweenthe sprung mass resonance frequency component and the unsprung massresonance frequency component is a frequency region in which, althoughthere is no sensation of reduced gravitational acceleration, there is asensation similar to the quick, frequent bouncing experienced by aperson on horseback when riding at a trot, that is, an up-and-downmotion followed by the entire body, and is referred to as the bounceregion (3-6 Hz). The region in which the unsprung mass resonancefrequency component is present is a frequency region in which, althoughvertical body mass movement is not experienced, frequent vibrations areconveyed to a part a passenger's body, such as the thighs, and isreferred to as a flutter region (6-23 Hz).

FIG. 12 is a control block diagram showing frequency-sensitive controlin sprung mass vibration damping control in the first embodiment. A bandelimination filter 350 cuts out noise apart from the frequency componentin the wheel speed sensor value that is used to perform control. Apredetermined frequency region splitter unit 351 splits the frequencycomponent into a float region, a bounce region, and a flutter region. AHilbert transform processor unit 352 performs a Hilbert transform uponthe split frequency bands, converting them to scalar quantities(specifically, areas calculated using amplitude and frequency band)based on the amplitude of the frequency. A vehicle vibrational systemweight-setting unit 353 sets weights for the vibration from the floatregion, bounce region, and flutter region frequency bands that isactually propagated to the vehicle. A human sensation weight-settingunit 354 sets weights for the vibration from the float region, bounceregion, and flutter region frequency bands that is actually propagatedto a passenger.

The setting of human sensation weights will now be described. FIG. 13 isa correlation graph showing human sense profiles plotted againstfrequency. As shown in FIG. 13, passenger sensitivity to frequencies iscomparatively low in the low-frequency float region, with sensitivitygradually increasing as one moves into regions of higher frequencies.Frequencies in the flutter region and higher-frequency regions becomeprogressively harder to transmit to the passenger. In view of this, thefloat region human sensation weight Wf is set at 0.17, the bounce regionhuman sensation weight Wh is set higher than Wf at 0.34, and the flutterregion human sensation weight Wb is set higher than Wf and Wh at 0.38.It is thereby possible to increase the correlation between the scalarquantities for the various frequency bands and the vibration actuallypropagated to a passenger. These two weighting factors may be altered,as appropriate, according to vehicle concept or passenger preferences.

A weight-determining device 355 calculates the proportions occupied bythe weight for each of the frequency bands. Defining a as the floatregion weight, b as the bounce region weight, and c as the flutterregion weight, the weighting factor for the float region is (a/(a+b+c)),the weighting factor for the bounce region is (b/(a+b+c)), and theweighting factor for the flutter region is (c/(a+b+c)).

A scalar quantity calculator unit 356 multiplies the scalar quantitiesfor the various frequency bands calculated by the Hilbert transformprocessor unit 352 by the device 355, and outputs final scalarquantities. The process up to this point is performed on the wheel speedsensor values for each of the wheels.

A maximum value-selecting unit 357 selects the maximum value out of thefinal scalar quantities calculated for each of the four wheels. The 0.01at the bottom is set to avoid having 0 as a denominator, as the total ofthe maximum values is used as a denominator in a subsequent process. Aproportion calculator unit 358 calculates proportions using the total ofthe maximum scalar quantity values for each of the frequency bands asthe denominator and the maximum scalar quantity value for the frequencyband equivalent to a float region as the numerator. In other words, theproportion of contamination (hereafter, simply “proportion”) in thefloat region contained in all vibration components is calculated. Asprung mass resonance filter 359 performs a filter process having asprung mass resonance frequency of roughly 1.2 Hz on the calculatedproportion, and extracts a sprung mass resonance frequency bandcomponent representing the float region from the calculated proportion.In other words, because the float region is present at roughly 1.2 Hz,it is believed that the proportion of this region will also vary around1.2 Hz. The finally extracted proportion is then outputted to thedamping force control unit 35, and a frequency-sensitive damping forcecontrol amount corresponding to the proportion is outputted.

FIG. 14 is a plot showing the relationship between the proportion ofvibration contamination and damping force in a float region in thefrequency-sensitive control of the first embodiment. As shown in FIG.14, a high damping force is set when the float region occupies a largeproportion, thereby reducing the vibration level of sprung massresonance. Even if a high damping force is set, no high-frequencyvibration or bounce-like vibration is transmitted to the passengerbecause the bounce region and flutter region occupy small proportions.Meanwhile, setting a low damping force when the float region occupies asmall proportion reduces the vibration transmission profile at and abovesprung mass resonance, minimizing high-frequency vibration and yieldinga smooth ride.

The benefits of frequency-sensitive control in a comparison offrequency-sensitive control and skyhook control will now be described.FIG. 15 is a wheel speed frequency profile detected by a wheel speedsensor 5 in certain driving conditions. This profile especially appearswhen driving on road surfaces having continuous small bumps, such ascobbled roads. When skyhook control is performed while driving on a roadsurface exhibiting this profile, the problem arises that, becausedamping force is determined using the peak amplitude value in skyhookcontrol, any degradation in phase estimation for high-frequencyvibrational input will cause an extremely high damping force to be setat the wrong timing, leading to exacerbated high-frequency vibration.

By contrast, if control is performed using scalar quantities rather thanvectors, as in frequency-sensitive control, the float region occupies asmall proportion on road surfaces such as that shown in FIG. 15, leadingto a low damping force being set. Thus, even if the amplitude of flutterregion vibration is high, the vibration transmission profile issufficiently reduced, allowing the exacerbation of high-frequencyvibration to be avoided. As shown by the foregoing, high-frequencyvibration can be minimized via scalar quantity-based frequency-sensitivecontrol in regions where control is difficult due to degradations inphase estimation precision even if skyhook control is performed using anexpensive sensor.

(Unsprung Mass Vibration Damping Control Unit)

Next, the configuration of the unsprung mass vibration damping controlunit will be described. As discussed in the context of the conventionalvehicle shown in FIG. 8B, a resonance frequency band is also present intires, as they possess both a modulus of elasticity and a dampingcoefficient. However, because a tire has a mass that is less than thatof the sprung mass, and a high modulus of elasticity as well, the bandis present toward frequencies higher than sprung mass resonance. Theunsprung mass resonance component causes tire rumbling in the unsprungmass, potentially degrading ground contact. In addition, rumbling in theunsprung mass can be uncomfortable for passengers. Thus, damping forceis set according to the unsprung mass resonance component in order tominimize unsprung mass resonance-induced rumbling.

FIG. 16 is a block diagram showing a control configuration for unsprungmass vibration damping control in the first embodiment. An unsprung massresonance component extractor unit 341 applies a band-pass filter to thewheel speed variation outputted from the deviation calculator unit 321 bof the driving state estimator unit 32 to extract an unsprung massresonance component. The unsprung mass resonance component is extractedfrom the region at roughly 10-20 Hz in the wheel speed frequencycomponent. An envelope waveform-forming unit 342 scalarizes theextracted unsprung mass resonance component, and forms an envelopewaveform using an envelope filter. A gain multiplier unit 343 multipliesthe scalarized unsprung mass resonance component by the gain, calculatesan unsprung mass vibration damping force control amount, which isoutputted to the damping force control unit 35. In the first embodiment,an unsprung mass resonance component is extracted by applying aband-pass filter to the wheel speed variation outputted from thedeviation calculator unit 321 b of the driving state estimator unit 32,but it is also acceptable to apply a band-pass filter to the valuedetected by the wheel speed sensor to extract the unsprung massresonance component, or for the driving state estimator unit 32 toestimate the unsprung mass speed along with the sprung mass speed toextract an unsprung mass resonance component.

(Configuration of Damping Force Control Unit)

Next, the configuration of the damping force control unit 35 will bedescribed. FIG. 17 is a control block diagram showing a controlconfiguration for a damping force control unit of the first embodiment.The driver input damping force control amount outputted from the driverinput control unit 31, the S/A orientation control amount outputted fromthe skyhook control unit 33 a, the frequency-sensitive damping forcecontrol amount outputted from the frequency-sensitive control unit 33 b,the unsprung mass vibration damping force control amount outputted fromthe unsprung mass vibration damping control unit 34, and the strokespeed calculated by the driving state estimator unit 32 are inputtedinto an equivalent viscous damping coefficient converter unit 35 a,which converts these values into an equivalent viscous dampingcoefficient.

A damping coefficient-reconciling unit 35 b reconciles which dampingcoefficient, out of the damping coefficients converted by the equivalentviscous damping coefficient converter unit 35 a (hereafter referred toindividually as the driver input damping coefficient k1, the S/Aorientation damping coefficient k2, the frequency-sensitive dampingcoefficient k3, and the unsprung mass vibration damping coefficient k4),is used to perform control, and outputs a final damping coefficient. Acontrol signal converter unit 35 c converts a control signal (commandedcurrent value) to be sent to the S/As 3 on the basis of the dampingcoefficient reconciled by the damping coefficient-reconciling unit 35 band the stroke speed, and outputs the signal to the S/As 3.

(Damping Coefficient-Reconciling Unit)

Next, the specifics of the reconciliation performed by the dampingcoefficient-reconciling unit 35 b will be described. The vehicle controldevice of the first embodiment has four control modes. The first mode isstandard mode, for situations in which suitable steering conditions areobtainable while driving on general urban roads. The second mode issports mode, for situations in which stable steering conditions areavailable while aggressively driving along winding roads and the like.The third mode is comfort mode, for situations in which priority isgiven to comfort while driving, such as when starting off at low vehiclespeeds. The fourth mode is highway mode, for situations involvingdriving at high vehicle speeds on highways and the like with multiplestraight sections.

In standard mode, priority is given to unsprung mass vibration dampingcontrol performed by the unsprung mass vibration damping control unit 34while skyhook control is being performed by the skyhook control unit 33a. In sports mode, skyhook control is performed by the skyhook controlunit 33 a and unsprung mass vibration damping control is performed bythe unsprung mass vibration damping control unit 34 while givingpriority to driver input control performed by the driver input controlunit 31. In comfort mode, priority is given to unsprung mass vibrationdamping control performed by the unsprung mass vibration damping controlunit 34 while frequency-sensitive control is being performed by thefrequency-sensitive control unit 33 b. In highway mode, the controlamount for the unsprung mass vibration damping control performed by theunsprung mass vibration damping control unit 34 is added to the skyhookcontrol performed by the skyhook control unit 33 a while given priorityto the driver input control performed by the driver input control unit31. Damping coefficient reconciliation in these various modes will nowbe described.

(Reconciliation in Standard Mode)

FIG. 18 is a flow chart of a damping coefficient reconciliation processperformed during a standard mode in the first embodiment.

In step S1, it is determined whether the S/A orientation dampingcoefficient k2 is greater than the unsprung mass vibration dampingcoefficient k4, and, if so, the process continues to step S4, and k2 isset as the damping coefficient.

In step S2, the scalar quantity proportion of the flutter region iscalculated on the basis of the scalar quantities for the float region,bounce region, and flutter region described in the context of thefrequency-sensitive control unit 33 b.

In step S3, it is determined whether the proportion of the flutterregion is equal to or greater than a predetermined value, and, if so,the process continues to step S4 and the low value k2 is set as thedamping coefficient for fear of high-frequency vibration reducing ridecomfort. On the other hand, if the proportion of the flutter region isless than the predetermined value, there is no worry of high-frequencyvibration reducing ride comfort even if a high damping coefficient isset, so the process continues to step S5, and k4 is set as thecoefficient.

In standard mode, as discussed above, priority is given, as a rule, tounsprung mass vibration damping control, which minimizes resonance inthe unsprung mass. However, if the damping force required for skyhookcontrol is less than the damping force required for unsprung massvibration damping control, and the flutter region occupies a largeproportion, the damping force for skyhook control is set so as to avoidexacerbating the high-frequency vibration profile in order to meet therequirements of unsprung mass vibration damping control. This allows anoptimal damping profile to be obtained according to the driving state,allowing high-frequency vibration-induced degradations of ride comfortto be avoided while simultaneously achieving a flat vehicle body feel.

(Reconciliation in Sports Mode)

FIG. 19 is a flow chart of a damping coefficient reconciliation processperformed during a sports mode in the first embodiment.

In step S11, the damping force distribution factors for the four wheelsare calculated on the basis of the driver input damping coefficients k1for the four wheels set during driver input control. Defining k1fr asthe front right wheel driver input damping coefficient, k1fl as thefront left wheel driver input damping coefficient, k1rr as the rearright wheel driver input damping coefficient, k1rl as the rear leftwheel driver input damping coefficient, and xfr, xfl, xrr, and xrl asthe damping force distribution factors for the different wheels, thedistribution factors are calculated as follows:

xfr=k1fr/(k1fr+k1fl+k1rl)

xfl=k1fl/(k1fr+k1fl+k1rl)

xrr=k1rr/(k1fr+k1fl+k1rl)

xrl=k1rl/(k1fr+k1fl+k1rl)

In step S12, it is determined whether a damping force distributionfactor x is within a predetermined range (greater than a and less thanβ), and, if so, distribution is determined to be roughly equal for allthe wheels, and the process continues to step S13; if even one factor isoutside the predetermined range, the process continues to step S16.

In step S13, it is determined whether the unsprung mass vibrationdamping coefficient k4 is greater than the driver input dampingcoefficient k1, and, if so, the process continues to step S15, and k4 isset as a first damping coefficient k. On the other hand, if the unsprungmass vibration damping coefficient k4 is equal to or less than thedriver input damping coefficient k1, the process continues to step S14,and k1 is set as the first damping coefficient k.

In step S16, it is determined whether the unsprung mass vibrationdamping coefficient k4 is the maximum value max that can be set for theS/As 3; if so, the process continues to step S17, and, if not, theprocess continues to step S18.

In step S17, the maximum value for the driver input damping coefficientsk1 for the four wheels is the unsprung mass vibration dampingcoefficient k4, and the damping coefficient that satisfies the dampingforce distribution factor is calculated as the first damping coefficientk. In other words, a value is calculated such that the dampingcoefficient is maximized while the damping force distribution factor issatisfied.

In step S18, a damping coefficient such that the damping forcedistribution factor is satisfied within a range in which the driverinput damping coefficients k1 for all four wheels are equal to orgreater than k4. In other words, a value is calculated such that thedamping force distribution factor set by the driver input control issatisfied, and the requirements of unsprung mass vibration dampingcontrol are also met.

In step S19, it is determined whether the first damping coefficients kset in the abovementioned steps are less than the S/A orientationdamping coefficient k2 set during skyhook control; if so, k2 is set andthe process continues to step S20 due to the damping coefficientrequired by skyhook control being larger. On the other hand, if k isequal to or greater than k2, k is set and the process continues to stepS21.

In sports mode, as discussed above, priority is given, as a rule, tounsprung mass vibration damping control, which minimizes resonance inthe unsprung mass. However, because the damping force distributionfactor required by driver input control is closely related to thevehicle body orientation, and is particularly deeply related to changesin driver line of view caused by roll mode, foremost priority is givento ensuring the damping force distribution factor, rather than thedamping coefficient required by driver input control itself. Formovement that causes changes in vehicle body orientation whilepreserving the damping force distribution factor, selecting skyhookcontrol via select high allows a stable vehicle body orientation to bemaintained.

(Reconciliation in Comfort Mode)

FIG. 20 is a flow chart of a damping coefficient reconciliation processperformed during a comfort mode in the first embodiment.

In step S30, it is determined whether the frequency-sensitive dampingcoefficient k3 is greater than the unsprung mass vibration dampingcoefficient k4, and, if so, the process continues to step S32 and thefrequency-sensitive damping coefficient k3 is set. On the other hand, ifthe frequency-sensitive damping coefficient k3 is determined to be equalto or less than the unsprung mass vibration damping coefficient k4, theprocess continues to step S32 and the unsprung mass vibration dampingcoefficient k4 is set.

In comfort mode, as discussed above, priority is given, as a rule, tounsprung mass resonance damping control, which minimizes resonance inthe unsprung mass. Because frequency-sensitive control is performed assprung mass vibration damping control to begin with, and an optimaldamping coefficient for the road surface conditions is set, control canbe performed while ensuring ride comfort, allowing sensationsinsufficient ground contact caused by rattling in the unsprung mass tobe avoided through unsprung mass vibration damping control. In comfortmode, as in standard mode, it is acceptable for the damping coefficientto be switched according to the proportion of flutter in the frequencyscalar quantity. This allows for a super comfort mode in which ridecomfort is even better ensured.

(Reconciliation in Highway Mode)

FIG. 21 is a flow chart of a damping coefficient reconciliation processperformed during a highway mode in the first embodiment. The samereconciliation process as in sports mode is performed from steps S11 toS18; thus, description thereof will be omitted. In step S40, the S/Aorientation damping coefficient k2 yielded by skyhook control is addedto the reconciled first damping coefficient k yielded by the process upto step S18 and outputted.

In highway mode, as discussed above, the sum of the reconciled firstdamping coefficient k and the S/A orientation damping coefficient k2 isused to reconcile the damping coefficient. This operation will now bedescribed with reference to the drawings. FIG. 22 is a time chartshowing changes in damping coefficient when driving on a hilly roadsurface and a bumpy road surface. For instance, if an attempt is made tominimize swaying motion in the vehicle body caused by the effects ofslight hills in the road surface when driving at high vehicle speeds viaskyhook control alone, it is necessary to detect slight variations inwheel speed, which requires that a comparatively high skyhook controlgain be set. In such cases, swaying motion can be minimized, but bumpsin the road surface can lead to the control gain being too great,creating the risk of excessive damping force control being performed.This gives rise to concerns of degraded ride comfort or vehicle bodyorientation.

By contrast, because the first damping coefficient k is constantly set,as in highway mode, a certain level of damping force can be constantlyensured, allowing swaying motion in the vehicle body to be minimizedeven if a low damping coefficient is used in skyhook control. Inaddition, because there is no need to increase the skyhook control gain,bumps in the road surface can be managed using a normal control gain.Moreover, because skyhook control is performed in a state in which thedamping coefficient k is set, a process of reducing the dampingcoefficient can be operated in a semi-active control region, unlike inthe case of a damping coefficient limit, ensuring a stable vehicleorientation during high-speed driving.

(Mode Selection Process)

Next, a mode selection process for selecting among the various drivingmodes described above will be described. FIG. 23 is a flow chart of adriving state-based mode selection process performed by a dampingcoefficient-reconciling unit of the first embodiment.

In step S50, it is determined whether the vehicle is driving straightahead based on the value from the steering angle sensor 7; if so, theprocess continues to step S51, and if the vehicle is determined to be ina state of turning, the process continues to step S54.

In step S51, it is determined whether the vehicle speed is equal to orgreater than a predetermined vehicle speed VSP1 indicating a state ofhigh vehicle speed on the basis of the value from the vehicle speedsensor 8, and, if so, the process continues to step S52 and standardmode is selected. On the other hand, if the speed is less than VSP1, theprocess continues to step S53 and comfort mode is selected.

In step S54, it is determined whether the vehicle speed is equal to orgreater than a predetermined vehicle speed VSP1 indicating a state ofhigh vehicle speed on the basis of the value from the vehicle speedsensor 8, and, if so, the process continues to step S55 and highway modeis selected. On the other hand, if the speed is less than VSP1, theprocess continues to step S56 and sports mode is selected.

That is, standard mode is selected when driving at a high vehicle speedwhen driving straight ahead, thereby making it possible to stabilize thevehicle body orientation via skyhook control, ensure ride comfort byminimizing high-frequency vibration-induced bouncing or fluttering, andminimizing resonance in the unsprung mass. Selecting comfort mode whendriving at low speeds makes it possible to minimize resonance in theunsprung mass while minimizing the transmission of vibration such asbouncing or fluttering to passengers.

Meanwhile, highway mode is selected when driving at a high vehicle speedin a state of turning, thereby performing control using a value to whicha damping coefficient has been added; thus, high damping force isyielded as a rule. It is thus possible to minimize unsprung massresonance while actively ensuring the unsprung mass resonance duringturning via driver input control, even when traveling at a high vehiclespeed. Selecting sports mode when driving at a low vehicle speed allowsunsprung mass resonance to be minimized while actively ensuring thevehicle body orientation during turning via driver input control andperforming skyhook control as appropriate, thereby allowing for drivingwith a stable vehicle orientation.

In the first embodiment, an example of a mode selection process in whichthe driving state is detected and the mode is automatically switched hasbeen presented, but it is also possible to provide a mode switch or thelike that can be operated by a driver to select the driving mode. Thisyields ride comfort and turning performance matching the driver'sdesired driving state.

As described above, the first embodiment yields the following effects.

(1) Provided are a driving state estimator unit 32 (a state quantitydetection device) for estimating pitch rate (a state quantity indicatingvehicle body orientation), and a brake orientation control amountcalculator unit 334 (a friction brake orientation control device) forminimizing pitching motion in the vehicle body orientation by applyingbraking torque from brakes 20 (friction brakes) at least to the frontwheels and minimizing bouncing motion in the vehicle body orientation byapplying braking torque from the brakes 20 to all four wheels, the brakeorientation control amount calculator unit 334 prioritizing minimizingpitching motion over minimizing bouncing motion.

Generally, the brakes 20 are capable of controlling both bounce andpitch; thus, it is preferable that they control both. However, whenbounce control is performed by the brakes 20, braking force is appliedto all four wheels simultaneously, and there is a strong sense ofdeceleration even in directions of low control priority despite thedifficulty in obtaining control effects, tending to create an unnaturalfeed for the driver. Thus, a configuration has been adopted in whichminimizing pitching motion over minimizing bouncing motion isprioritized for the brakes 20, thereby specializing the first embodimentin particular for pitch control. This allows sensations of decelerationto be minimized, thereby reducing the level of unnatural sensationsexperienced by a passenger.

In the first embodiment, if braking force is applied when the pitch rateVp is positive, i.e., the front wheel side of the vehicle is lowered,the front wheel side will sink further lower, augmenting pitch motion;thus, braking force is not applied in such cases. On the other hand,when the pitch rate Vp is negative, i.e., the front wheel side of thevehicle is raised, the braking pitch moment will impart braking force,minimizing the rising of the front wheel side. This ensures the driver'sfield of view and makes the area ahead easier to see, contributing toimproved senses of safety and flatness of ride. Because braking torqueis generated only when the front side of the vehicle body is raised, theamount of deceleration generated can be reduced compared to arrangementsin which braking torque is generated both when the front side is raisedand when it is lowered. In addition, the actuators need only be operatedat half the frequency as usual, allowing inexpensive actuators to beused.

An example specialized for pitch control has been presented for thefirst embodiment, but a configuration in which minimizing pitchingmotion is prioritized in the course of controlling both pitching motionand bouncing motion, or one in which the bouncing motion control amountis multiplied by a gain that will reduce the amount of controlperformed, is also acceptable. The object of the present invention willbe fulfilled as long as pitch control is prioritized over bouncecontrol.

An example in which skyhook control is applied as pitch control has beenpresented for the first embodiment, but other control logics arepossible as long as pitch rate-minimizing braking torque is outputted.

(2) The brake orientation control amount calculator unit 334 controlsbraking torque so that the rate of change in vehicle body decelerationis equal to or less than a predetermined value.

Specifically, a jerk threshold value limiter unit 3342 d determineswhether the rate of change in the calculated target deceleration, i.e.,jerk, is within preset deceleration jerk threshold value and releasejerk threshold value ranges, and whether the target deceleration iswithin a forward/reverse acceleration limit value range. If any of thethreshold values is exceeded, the target deceleration is corrected to avalue within the ranges for the jerk threshold values. If the targetdeceleration exceeds the limit value, it is set to within the limitvalue. It is thereby possible to generate deceleration so as not tocreate an unnatural feel for the driver.

(3) The driving state estimator unit 32 detects vehicle body orientationon the basis of the values from the wheel speed sensors 5 for each ofthe wheels. All states are thus estimated using the wheel speed sensors5 with which all vehicles are generally equipped, without the need forexpensive sensors such as a sprung weight vertical acceleration sensoror a stroke sensor, thereby allowing costs and the number of parts to bereduced, and the ease with which components are installed in the vehicleto be improved.

(4) The driving state estimator unit 32 (a driving state detectiondevice) estimates the driving state on the basis of an estimable activeskyhook model regardless of the attributes of the sprung mass speed andthe stroke speed.

This stabilizes filter response and allows suitable estimation precisionto be obtained. It should be noted that, even if an active skyhook modelis adopted, only semi-active control is actually possible; thus, thecontrollable range is halved. The estimated sprung mass speed istherefore less than the actual speed in the frequency band from sprungmass resonance down. However, it is the phase that is most vital in thecontext of skyhook control, and skyhook control can be achieved as longas the correspondence between phase and attribute can be maintained;this is unproblematic because the sprung mass speed can be adjustedusing the other coefficients or the like.

(5) The driving state estimator unit 32 (a driving state detectiondevice) estimates the driving state using a four-wheel model constructedusing a bounce term representing the four-wheel vertical movement, apitch term representing front and rear wheel vertical movement, a rollterm representing left and right wheel vertical movement, and a warpterm representing diagonal wheel vertical movement.

In other words, one corresponding component is lacking when the sprungmass speed for all four wheels is divided into roll term, pitch term,and bounce term modes, destabilizing the solution. Thus, the warp termrepresenting the movement of diagonal wheels is introduced in order toallow the abovementioned terms to be estimated.

(6) Provided are a driving state estimator unit 32 (sensor) fordetecting pitch rate (a state quantity indicating vehicle bodyorientation), and a brake orientation control amount calculator unit 334(controller) for prioritizing minimizing pitching motion over minimizingbouncing motion in the course of performing friction brake orientationcontrol capable of minimizing pitching motion and bouncing motion in thevehicle body orientation using braking torque from the friction brakes.

As a result, sensations of deceleration can be minimized, reducing thelevel of unnatural sensations experienced by a passenger. An example inwhich skyhook control is applied as pitch control has been presented forthe first embodiment, but other control logics are possible as long aspitch rate-minimizing braking torque is outputted.

(7) A driving state estimator unit 32 (sensor) for detecting pitch rate(a state quantity indicating vehicle body orientation) is comprised,and, in the course of performing friction brake orientation controlcapable of minimizing pitching motion and bouncing motion in the vehiclebody orientation using braking torque from the friction brakes, thebrake orientation control amount calculator unit 334 prioritizesminimizing the pitching motion over minimizing the bouncing motion.

As a result, sensations of deceleration can be minimized, reducing thelevel of unnatural sensations experienced by a passenger. An example inwhich skyhook control is applied as pitch control has been presented forthe first embodiment, but other control logics are possible as long aspitch rate-minimizing braking torque is outputted.

1. A vehicle control device comprising: a state quantity detectiondevice configured to detect a state quantity indicating a vehicle bodyorientation; and a friction brake orientation control device configuredto minimize pitching motion in the vehicle body orientation by applyingbraking torque from a friction brake at least to a front wheel, andminimize bouncing motion in the vehicle body orientation by applyingbraking torque from the friction brake to four wheels, the frictionbrake orientation control device being configured to prioritizeminimizing the pitching motion over minimizing the bouncing motion. 2.The vehicle control device according to claim 1, wherein the frictionbrake orientation control is configured to control the braking torque sothat changes in the rate of vehicle body deceleration is equal to orless than a predetermined value.
 3. The vehicle control device accordingto claim 1, wherein the state quantity detection device is configured todetect the vehicle body orientation on the basis of values from wheelspeed sensors for the four wheels.
 4. The vehicle control deviceaccording to claim 1, wherein the driving state detection device isconfigured to estimate driving state on the basis of an estimable activeskyhook model when a sprung mass speed and a stroke speed have positiveand negative attributes.
 5. The vehicle control device according toclaim 1, wherein the driving state detection device is configured toestimate driving state using a four-wheel model constructed using abounce term representing four-wheel vertical movement, a pitch termrepresenting front and rear wheel vertical movement, a roll termrepresenting left and right wheel vertical movement, and a warp termrepresenting diagonal wheel vertical movement.
 6. A vehicle controldevice comprising: a sensor configured to detect a state quantityindicating a vehicle body orientation; and a controller configured tominimize pitching motion in the vehicle body orientation by applyingbraking torque from a friction brake to at least a front wheel, andapplying braking torque from a friction brake to four wheels toprioritize minimizing pitching motion over minimizing bouncing motionwhen performing friction brake orientation control for minimizingbouncing motion in the vehicle body orientation.
 7. A vehicle controlmethod comprising: detecting, using a sensor, a state quantityindicating a vehicle body orientation; and minimizing, using acontroller, pitching motion in the vehicle body orientation by applyingbraking torque from a friction brake to at least a front wheel, andapplying braking torque from the friction brake to four wheels toprioritize minimizing pitching motion over minimizing bouncing motionwhen performing friction brake orientation control for minimizingbouncing motion in the vehicle body orientation.
 8. The vehicle controldevice according to claim 2, wherein the state quantity detection deviceis configured to detect the vehicle body orientation on the basis ofvalues from wheel speed sensors for the four wheels.
 9. The vehiclecontrol device according to claim 2, wherein the driving state detectiondevice is configured to estimate driving state on the basis of anestimable active skyhook model when a sprung mass speed and a strokespeed have positive and negative attributes.
 10. The vehicle controldevice according to claim 3, wherein the driving state detection deviceis configured to estimate driving state on the basis of an estimableactive skyhook model when a sprung mass speed and a stroke speed havepositive and negative attributes.
 11. The vehicle control deviceaccording to claim 2, wherein the driving state detection device isconfigured to estimate driving state using a four-wheel modelconstructed using a bounce term representing four-wheel verticalmovement, a pitch term representing front and rear wheel verticalmovement, a roll term representing left and right wheel verticalmovement, and a warp term representing diagonal wheel vertical movement.12. The vehicle control device according to claim 3, wherein the drivingstate detection device is configured to estimate driving state using afour-wheel model constructed using a bounce term representing four-wheelvertical movement, a pitch term representing front and rear wheelvertical movement, a roll term representing left and right wheelvertical movement, and a warp term representing diagonal wheel verticalmovement.
 13. The vehicle control device according to claim 4, whereinthe driving state detection device is configured to estimate drivingstate using a four-wheel model constructed using a bounce termrepresenting four-wheel vertical movement, a pitch term representingfront and rear wheel vertical movement, a roll term representing leftand right wheel vertical movement, and a warp term representing diagonalwheel vertical movement.